Controller of internal combustion engine

ABSTRACT

A device for controlling an internal combustion engine, comprising a variable valve mechanism for varying opening areas (valve lift) or the working angles (valve-opening periods) of at least either the intake valves or the exhaust valves, wherein a pressure in the cylinder is calculated based on the opening area or the working angle of at least either the intake valve or the exhaust valve varied by the variable valve mechanism, and the internal combustion engine is controlled based on the pressure in the cylinder. Upon calculating the pressure in the cylinder based on the opening areas or the working angles of the intake and exhaust valves, it is possible to more suitably control the internal combustion engine based not only upon the peak combustion pressure in the cylinder like when a combustion pressure sensor is used but also upon a pressure in the cylinder at a moment other than the peak combustion pressure.

This is a Divisional of application Ser. No. 11/131,404 filed May 18,2005, which is a Divisional of application Ser. No. 10/450,152 filedJun. 11, 2003 (now U.S. Pat. No. 6,948,478, issued Sep. 27, 2005), whichin turn is a National Stage of PCT/JP01/10917 filed Dec. 12, 2001. Theentire disclosure of the prior applications is hereby incorporated byreference herein in its entirety.

TECHNICAL FIELD

The present invention relates to a device for controlling internalcombustion engines.

BACKGROUND ART

There have heretofore been known devices for controlling internalcombustion engines based upon the pressure in the cylinders. JapaneseUnexamined Patent Publication (Kokai) No. 9-53503 discloses a device ofthis kind for controlling internal combustion engines. In the device forcontrolling internal combustion engines disclosed in the above JapaneseUnexamined Patent Publication (Kokai) No. 9-53503, the amount of fuelinjection and the ignition timing are calculated based on an outputvalue of a cylinder pressure sensor that detects the pressure in thecylinder.

When the pressure in the cylinder is detected by the cylinder pressuresensor as in the device for controlling internal combustion enginesdisclosed in Japanese Unexamined Patent Publication (Kokai) No. 9-53503,however, the pressure in the cylinder that is detected is a peakcombustion pressure at a moment of a crank angle of 10 to 15 degreesafter the top dead center in the compression stroke. With the device forcontrolling internal combustion engines disclosed in Japanese UnexaminedPatent Publication (Kokai) No. 9-53503, therefore, it is not possible tocontrol the internal combustion engines based on the pressure in thecylinder other than the peak combustion pressure in the cylinder, suchas the pressue at the top dead center in the compression stroke. Inaddition, with the device for controlling internal combustion engines,which does not consider a change in the pressure in the cylinderaccompanying a change in the opening areas of the intake and exhaustvalves or a change in the working angle as a result of employing avariable valve mechanism as taught in Japanese Unexamined PatentPublication (Kokai) No. 9-53503, it is not possible to suitably controlthe internal combustion engine when the opening areas of the intake andexhaust valves vary or when the working angles thereof vary.

There has further been known a device for controlling internalcombustion engines based on the temperature of a certain portion in thecylinder. The device for controlling internal combustion engines of thistype has been disclosed in, for example, Japanese Unexamined PatentPublication (Kokai) No. 4-81574. In the device for controlling internalcombustion engines taught in Japanese Unexamined Patent Publication(Kokai) No. 4-81574, the ignition timing is calculated based upon anoutput value of a wall temperature sensor that detects the temperatureof the cylinder wall, and the internal combustion engine is controlledbased on the thus calculated ignition timing.

In the device for controlling internal combustion engines taught in theabove Japanese Unexamined Patent Publication (Kokai) No. 4-81574, theignition timing is calculated based on the temperature of the cylinderwall, and the internal combustion engine is controlled based on theignition timing. However, there is a considerable difference between thetemperature of the cylinder wall and the temperature of gas in thecylinder. In addition, it can be said that a suitable ignition timing isdetermined based on the temperature of gas in the cylinder rather thanthe temperature of the cylinder wall. Therefore, the internal combustionengine cannot be suitably controlled by the device for controllinginternal combustion engines, which calculates the ignition timing basedon the temperature of the cylinder wall as taught in Japanese UnexaminedPatent Publication (Kokai) No. 4-81574. Moreover, with the device forcontrolling internal combustion engines, which does not consider achange in the temperature of gas in the cylinder accompanying a changein the opening areas of the intake and exhaust valves or a change in theworking angles thereof as a result of employing a variable valvemechanism taught in Japanese Unexamined Patent Publication (Kokai) No.4-81574, it is not possible to suitably control the internal combustionengine when the opening areas of the intake and exhaust valves vary orwhen the working angles thereof vary.

There has further been known a device for controlling internalcombustion engines based upon the ratio or amount of an internal EGR gasby being provided with a variable valve mechanism for at least theintake valves or the exhaust valves. As a device for controllinginternal combustion engines of this kind, there has been known the onedisclosed in, for example, Japanese Unexamined Patent Publication(Kokai) No. 9-209895. The device for controlling internal combustionengines disclosed in Japanese Unexamined Patent Publication (Kokai) No.9-209895 is equipped with a variable valve mechanism for shifting theopening/closing timing (valve timing) without varying the length ofperiod for opening the intake valve, and calculates the ignition timingbased on the sum of the amount of the internal EGR gas (recirculatingamount of the internal exhaust gas) or the amount of the burnt gas takeninto the cylinder again after blown back into the intake pipe and theamount of the burnt gas remaining in the cylinder, i.e. not exhaustedfrom the cylinder, thereby to control the internal combustion enginebased on the thus calculated ignition timing.

However, the device for controlling internal combustion enginesdisclosed in Japanese Unexamined Patent Publication (Kokai) No. 9-209895is not considering the amount of varying the opening areas of the intakeand exhaust valves, though it is considering the amount of shifting theopening/closing timing of the intake and exhaust valves at the time ofcalculating the amount of the internal EGR gas. When the variable valvemechanism is provided with a function for varying the valve-liftingamount and when the opening areas of the intake and exhaust valves arevaried by changing the valve-lifting amount, the real amount of theinternal EGR gas varies to a considerable degree even though theopening/closing timings of the intake and exhaust valves are notshifted. When the amount of the internal EGR gas is calculated by thedevice for controlling internal combustion engines disclosed in JapaneseUnexamined Patent Publication (Kokai) No. 9-209895 without considering achange in the opening areas of the intake and exhaust valves despite theopening areas of the intake and exhaust valves are changing, therefore,the calculated amount of the internal EGR gas becomes considerablydifferent from the real amount of the internal EGR gas. Namely, when theopening areas of the intake and exhaust valves are subject to change,the amount of the internal EGR gas is not correctly calculated by thedevice for controlling internal combustion engines taught in JapaneseUnexamined Patent Publication (Kokai) No. 9-209895, which considers achange in the amount of the internal EGR gas accompanying a change inthe opening areas of the intake and exhaust valves due, for example, tothe variable valve mechanism. Accordingly, it is not possible tosuitably control the internal combustion engines.

Further, the device for controlling the internal combustion enginesdisclosed in Japanese Unexamined Patent Publication (Kokai) No. 9-209895considers the amount of shifting the opening/closing timings of theintake and exhaust valves at the time of calculating the amount of theinternal EGR gas but does not consider the amount of changing therotational angle of the cam shafts that correspond to the periods ofopening the intake and exhaust valves. On the other hand, when thevariable valve mechanism is provided with a function for varying theworking angles of the intake and exhaust valves, i.e., with a functionfor increasing or decreasing the periods for opening the intake andexhaust valves and when the working angles of the intake and exhaustvalves are varied, the real amount of the internal EGR gas changes to aconsiderable degree even when the opening/closing timings of the intakeand exhaust valves are not shifted, i.e., even when the peak timing ofthe valve-lifting amount is not changed. When the amount of the internalEGR gas is calculated by the device for controlling internal combustionengines disclosed in Japanese Unexamined Patent Publication (Kokai) No.9-209895 without considering a change in the working angles of theintake and exhaust valves though the working angles of the intake andexhaust valves are changing, therefore, the calculated amount of theinternal EGR gas becomes considerably different from the real amount ofthe internal EGR gas. Namely, when the working angles of the intake andexhaust valves are subject to change, the amount of the internal EGR gasis not correctly calculated by the device for controlling internalcombustion engines taught in Japanese Unexamined Patent Publication(Kokai) No. 9-209895, which does not consider a change in the amount ofthe internal EGR gas accompanying a change in the working angles of theintake and exhaust valves due, for example, to the variable valvemechanism. Accordingly, it is not possible to suitably control theinternal combustion engines.

There has further been known a device for controlling internalcombustion engines equipped with a variable valve mechanism for at leasteither the intake valves or the exhaust valves, based on a degree ofturbulence in the cylinder that is estimated relying upon the openingarea of the intake valve varied by the variable valve mechanism. Adevice for controlling internal combustion engines of this type has beendisclosed in, for example, Japanese Unexamined Patent Publication(Kokai) No. 2000-73800. In the device for controlling internalcombustion engines disclosed in Japanese Unexamined Patent Publication(Kokai) No. 2000-73800, it is estimated that the degree of turbulence inthe cylinder decreases with a decrease in the opening area of the intakevalve that is varied by the variable valve mechanism.

Namely, in the device for controlling internal combustion enginesdisclosed in Japanese Unexamined Patent Publication (Kokai) No.2000-73800, it is estimated that the degree of turbulence in thecylinder decreases with a decrease in the opening area of the intakevalve that is varied by the variable valve mechanism. In practice,however, the degree of turbulence in the cylinder increases with adecrease in the opening area of the intake valve that is varied by thevariable valve mechanism. Therefore, if it is estimated that the degreeof turbulence in the cylinder decreases with a decrease in the openingarea of the intake valve like in the device for controlling internalcombustion engines as disclosed in Japanese Unexamined PatentPublication (Kokai) No. 2000-73800 and if the internal combustion engineis controlled based on the estimated degree of turbulence in thecylinder, then, the internal combustion engine is not suitablycontrolled when the opening area of the intake valve is varied by thevariable valve mechanism.

DISCLOSURE OF THE INVENTION

In view of the above problems, it is an object of the present inventionto provide a device for controlling an internal combustion engine basednot only upon a peak combustion pressure in the cylinder but also upon apressure in the cylinder at a moment other than the peak combustionpressure, thereby to suitably control the internal combustion engineeven when the opening areas or the working angles of the intake andexhaust valves are varied.

It is another object of the present invention to provide a device formore suitably controlling an internal combustion engine than when theinternal combustion engine is controlled based on the temperature of thecylinder wall even when the opening areas or the working angles of theintake and exhaust valves are varied.

It is a further object of the present invention to provide a device forsuitably controlling an internal combustion engine by correctlycalculating an amount of the internal EGR gas even when the openingareas or the working angles of the intake and exhaust valves are varied.

It is a further object of the present invention to provide a device forsuitably controlling an internal combustion engine by correctlyestimating a degree of turbulence in the cylinder even when the openingareas or the working angles of the intake and exhaust valves are variedby the variable valve mechanism.

According to a first aspect of the present invention, there is provideda device for controlling an internal combustion engine based on apressure in the cylinder, comprising a variable valve mechanism forvarying opening areas of at least either the intake valves or theexhaust valves, wherein a pressure in the cylinder is calculated basedon the opening area of at least either the intake valve or the exhaustvalve varied by the variable valve mechanism, and the internalcombustion engine is controlled based on the pressure in the cylinder.

Namely, in the device for controlling an internal combustion engineaccording to the first aspect of the invention, a pressure in thecylinder is calculated based on the opening area of at least either theintake valve or the exhaust valve varied by the variable valvemechanism. It is therefore possible to control the internal combustionengine based not only upon a peak combustion pressure in the cylinderbut also upon a pressure in the cylinder at a moment other than the peakcombustion pressure, unlike the case of detecting the pressure in thecylinder by the cylinder pressure sensor employed by the device forcontrolling internal combustion engines taught in Japanese UnexaminedPatent Publication (Kokai) No. 9-53503. Further, since the internalcombustion engine is controlled based on the pressure in the cylindercalculated relying upon the opening areas of at least either the intakevalves or the exhaust valves, it is possible to suitably control theinternal combustion engine even when the opening areas of the intake andexhaust valves are varied. More specifically, the pressure in thecylinder increases with an increase in the opening area of the intakevalve, the pressure in the cylinder being calculated based on theopening area of the intake valve, and the internal combustion engine isso controlled that the ignition timing is delayed with an increase inthe pressure in the cylinder. Alternatively, the pressure in thecylinder increases with an increase in the opening area of the intakevalve, the pressure in the cylinder being calculated based on theopening area of the intake valve, and the internal combustion engine isso controlled that the amount of fuel injection is increased with anincrease in the pressure in the cylinder.

According to a second aspect of the present invention, there is provideda device for controlling an internal combustion engine based on apressure in the cylinder, comprising a variable valve mechanism forvarying working angles of at least either the intake valves or theexhaust valves, wherein a pressure in the cylinder is calculated basedon the working angle of at least either the intake valve or the exhaustvalve varied by the variable valve mechanism, and the internalcombustion engine is controlled based on the pressure in the cylinder.

Namely, in the device for controlling an internal combustion engineaccording to the second aspect of the invention, a pressure in thecylinder is calculated based on the working angle of at least either theintake valve or the exhaust valve varied by the variable valvemechanism. It is therefore possible to control the internal combustionengine based not only upon a peak combustion pressure in the cylinderbut also upon a pressure in the cylinder at a moment other than the peakcombustion pressure, unlike the case of detecting the pressure in thecylinder by the cylinder pressure sensor employed by the device forcontrolling internal combustion engines taught in Japanese UnexaminedPatent Publication (Kokai) No. 9-53503. Further, since the internalcombustion engine is controlled based on the pressure in the cylindercalculated relying upon the working angles of at least either the intakevalves or the exhaust valves, it is possible to suitably control theinternal combustion engine even when the working angles of the intakeand exhaust valves are varied. More specifically, when the intake valveis fully closed after the bottom dead center in the intake stroke, thepressure in the cylinder increases with a decrease in the working angleof the intake valve, the pressure in the cylinder being calculated basedon the working angle of the intake valve, and the internal combustionengine is so controlled that the ignition timing is delayed with anincrease in the pressure in the cylinder. Further, when the intake valveis fully closed before the bottom dead center in the intake stroke, thepressure in the cylinder increases with an increase in the working angleof the intake valve, the pressure in the cylinder being calculated basedon the working angle of the intake valve, and the internal combustionengine is so controlled that the ignition timing is delayed with anincrease in the pressure in the cylinder. Alternatively, the pressure inthe cylinder increases with a decrease in the working angle of theintake valve, the pressure in the cylinder being calculated based on theworking angle of the intake valve, and the internal combustion engine isso controlled that the amount of fuel injection is increased with anincrease in the pressure in the cylinder.

According to a third aspect of the present invention, there is provideda device for controlling an internal combustion engine, comprising avariable valve mechanism for varying opening areas and working angles ofat least either the intake valves or the exhaust valves, wherein apressure in the cylinder is calculated based on the opening area and theworking angle of at least either the intake valve or the exhaust valvevaried by the variable valve mechanism, and the internal combustionengine is controlled based on the pressure in the cylinder.

Namely, in the device for controlling an internal combustion engineaccording to the third aspect of the invention, a pressure in thecylinder is calculated based on the opening area and the working angleof at least either the intake valve or the exhaust valve varied by thevariable valve mechanism, and the internal combustion engine iscontrolled based upon the pressure in the cylinder. It is thereforepossible to suitably control the internal combustion engine by morecorrectly calculating the pressure in the cylinder than a case ofcalculating the pressure in the cylinder based upon the opening areasonly of the intake and exhaust valves but not upon the working angles ofthe intake and exhaust valves, or than a case of calculating thepressure in the cylinder based upon the working angles only of theintake and exhaust valves but not upon the opening areas of the intakeand exhaust valves.

According to a fourth aspect of the present invention, there is provideda device for controlling an internal combustion engine based on atemperature of a certain portion in the cylinder, comprising a variablevalve mechanism for varying opening areas of at least either the intakevalves or the exhaust valves, wherein a temperature of gas in thecylinder is calculated based on the opening area of at least either theintake valve or the exhaust valve varied by the variable valvemechanism, and the internal combustion engine is controlled based on thetemperature of gas in the cylinder.

In the device for controlling an internal combustion engine according tothe fourth aspect of the invention, a temperature of gas in the cylinderis calculated based on the opening area of at least either the intakevalve or the exhaust valve varied by the variable valve mechanism, andthe internal combustion engine is controlled based upon the temperatureof gas in the cylinder. It is therefore possible to more suitablycontrol the internal combustion engine than when the internal combustionengine is controlled based on the temperature of the cylinder wall thatis done by the device for controlling internal combustion engines taughtin Japanese Unexamined Patent Publication (Kokai) No. 4-81574. Further,since the internal combustion engine is controlled based on thetemperature of gas in the cylinder calculated relying upon the openingarea of at least either the intake valves or the exhaust valves, it ispossible to suitably control the internal combustion engine even whenthe opening areas of the intake and exhaust valves are varied. Morespecifically, the temperature of gas in the cylinder increases with anincrease in the opening area of the intake valve, the temperature of gasin the cylinder being calculated based on the opening area of the intakevalve, and the internal combustion engine is so controlled that theignition timing is delayed with an increase in the temperature of gas inthe cylinder.

According to a fifth aspect of the present invention, there is provideda device for controlling an internal combustion engine based on atemperature of a certain portion in the cylinder, comprising a variablevalve mechanism for varying working angles of at least either the intakevalves or the exhaust valves, wherein a temperature of gas in thecylinder is calculated based on the working angle of at least either theintake valve or the exhaust valve varied by the variable valvemechanism, and the internal combustion engine is controlled based on thetemperature of gas in the cylinder.

In the device for controlling an internal combustion engine according tothe fifth aspect of the invention, the temperature of gas in thecylinder is calculated based on the working angle of at least either theintake valve or the exhaust valve varied by the variable valvemechanism, and the internal combustion engine is controlled based uponthe temperature of gas in the cylinder. It is therefore possible tosuitably control the internal combustion engine better than when theinternal combustion engine is controlled based on the temperature of thecylinder wall, as done by the device for controlling internal combustionengines taught in Japanese Unexamined Patent Publication (Kokai) No.4-81574. Further, since the internal combustion engine is controlledbased on the temperature of gas in the cylinder calculated relying uponthe working angles of at least either the intake valves or the exhaustvalves, it is possible to suitably control the internal combustionengine even when the working angles of the intake and exhaust valves arevaried. More specifically, when the intake valve is fully closed afterthe bottom dead center in the intake stroke, the temperature of gas inthe cylinder increases with an increase in the working angle of theintake valve, the temperature of gas in the cylinder being calculatedbased on the working angle of the intake valve, and the internalcombustion engine is so controlled that the ignition timing is delayedwith an increase in the temperature of gas in the cylinder. Further,when the intake valve is fully closed before the bottom dead center inthe intake stroke, the temperature of gas in the cylinder increases witha decrease in the working angle of the intake valve, the temperature ofgas in the cylinder being calculated based on the working angle of theintake valve, and the internal combustion engine is so controlled thatthe ignition timing is delayed with an increase in the temperature ofgas in the cylinder.

According to a sixth aspect of the present invention, there is provideda device for controlling an internal combustion engine, comprising avariable valve mechanism for varying opening areas and working angles ofat least either the intake valves or the exhaust valves, wherein atemperature of gas in the cylinder is calculated based on the openingarea and the working angle of at least either the intake valve or theexhaust valve varied by the variable valve mechanism, and the internalcombustion engine is controlled based on the temperature of gas in thecylinder.

Namely, in the device for controlling an internal combustion engineaccording to the sixth aspect of the invention, a temperature of gas inthe cylinder is calculated based on the opening area and the workingangle of at least either the intake valve or the exhaust valve varied bythe variable valve mechanism, and the internal combustion engine iscontrolled based upon the temperature of gas in the cylinder. It istherefore possible to more suitably control the internal combustionengine by correctly calculating the temperature of gas in the cylinderthan a case of calculating the temperature of gas in the cylinder basedupon the opening areas only of the intake and exhaust valves but notupon the working angles of the intake and exhaust valves, or than a caseof calculating the temperature of gas in the cylinder based upon theworking angles only of the intake and exhaust valves but not upon theopening areas of the intake and exhaust valves.

According to a seventh aspect of the present invention, there isprovided a device for controlling an internal combustion engine,comprising a variable valve mechanism for at least either the intakevalves or the exhaust valves thereby to control the internal combustionengine based on a ratio or amount of the internal EGR gas, wherein aratio or amount of the internal EGR gas is calculated based on theopening area of at least either the intake valve or the exhaust valvevaried by the variable valve mechanism, and the internal combustionengine is controlled based on the ratio or amount of the internal EGRgas.

In the device for controlling an internal combustion engine according tothe seventh aspect of the invention, a ratio or amount of the internalEGR gas is calculated based on the opening area of at least either theintake valve or the exhaust valve varied by the variable valvemechanism, and the internal combustion engine is controlled based uponthe ratio or amount of the internal EGR gas. It is therefore possible tosuitably control the internal combustion engine by correctly calculatingthe ratio or amount of the internal EGR gas without considering a changein the opening areas of the intake and exhaust valves effected by thevariable valve mechanism unlike that of the device for controllinginternal combustion engines taught in Japanese Unexamined PatentPublication (Kokai) No. 9-209895. Namely, even when the opening areas ofthe intake and exhaust valves are varied, it is possible to correctlycalculate the amount of the internal EGR gas and to suitably control theinternal combustion engine. More specifically, the ratio or amount ofthe internal EGR gas increases with an increase in the opening area ofthe intake valve, the ratio or amount of the internal EGR gas beingcalculated based on the opening area of the intake valve, and theinternal combustion engine is so controlled that the ignition timing isadvanced with an increase in the ratio or amount of the internal EGRgas.

According to an eighth aspect of the present invention, there isprovided a device for controlling an internal combustion engine,comprising a variable valve mechanism for at least either the intakevalves or the exhaust valves thereby to control the internal combustionengine based on a ratio or amount of the internal EGR gas, wherein aratio or amount of the internal EGR gas is calculated based on theworking angle of at least either the intake valve or the exhaust valvevaried by the variable valve mechanism, and the internal combustionengine is controlled based on the ratio or amount of the internal EGRgas.

In the device for controlling an internal combustion engine according tothe eighth aspect of the invention, a ratio or amount of the internalEGR gas is calculated based on the working angle of at least either theintake valve or the exhaust valve varied by the variable valvemechanism, and the internal combustion engine is controlled based uponthe ratio or amount of the internal EGR gas. It is therefore possible tosuitably control the internal combustion engine by correctly calculatingthe ratio or amount of the internal EGR gas without considering a changein the working angles of the intake and exhaust valves effected by thevariable valve mechanism unlike that of the device for controllinginternal combustion engines taught in Japanese Unexamined PatentPublication (Kokai) No. 9-209895. Namely, even when the working anglesof the intake and exhaust valves are varied, it is possible to correctlycalculate the amount of the internal EGR gas and to suitably control theinternal combustion engine. More specifically, the ratio or amount ofthe internal EGR gas increases with an increase in the working angle ofthe intake valve, the ratio or amount of the internal EGR gas beingcalculated based on the working angle of the intake valve, and theinternal combustion engine is so controlled that the ignition timing isadvanced with an increase in the ratio or amount of the internal EGRgas.

According to a ninth aspect of the present invention, there is provideda device for controlling an internal combustion engine, wherein a ratioor amount of the internal EGR gas is calculated based on the openingareas and the working angles of at least either the intake valves or theexhaust valves varied by a variable valve mechanism, and the internalcombustion engine is controlled based on the ratio or amount of theinternal EGR gas.

In the device for controlling an internal combustion engine according tothe ninth aspect of the invention, a ratio or amount of the internal EGRgas is calculated based on the opening area and the working angle of atleast either the intake valve or the exhaust valve varied by thevariable valve mechanism, and the internal combustion engine iscontrolled based upon the ratio or amount of the internal EGR gas. It istherefore possible to more suitably control the internal combustionengine by correctly calculating the ratio or amount of the internal EGRgas than a case of calculating the ratio or amount of the internal EGRgas based upon the opening areas only of the intake and exhaust valvesbut not upon the working angles of the intake and exhaust valves, orthan a case of calculating the ratio or amount of the internal EGR gasbased upon the working angles only of the intake and exhaust valves butnot upon the opening areas of the intake and exhaust valves.

According to a tenth aspect of the present invention, there is provideda device for controlling an internal combustion engine, comprising avariable valve mechanism for at least either the intake valves or theexhaust valves thereby to control the internal combustion engine basedon a degree of turbulence in the cylinder that is estimated based on theopening area of the intake valve varied by the variable valve mechanism,wherein it is so estimated that a degree of turbulence in the cylinderincreases with a decrease in the opening area of the intake valve variedby the variable valve mechanism, and the internal combustion engine iscontrolled based on the estimated degree of turbulence in the cylinder.

In the device for controlling an internal combustion engine according tothe tenth aspect of the invention, it is estimated that a degree ofturbulence in the cylinder increases with a decrease in the opening areaof the intake valve varied by the variable valve mechanism, and theinternal combustion engine is controlled based upon the estimated degreeof turbulence in the cylinder. Even when the opening area of the intakevalve is varied by the variable valve mechanism, therefore, the degreeof turbulence in the cylinder is correctly estimated and the internalcombustion engine is suitably controlled unlike that of using the devicefor controlling internal combustion engines taught in JapaneseUnexamined Patent Publication (Kokai) No. 2000-73800 according to whichit is so estimated that a degree of turbulence in the cylinder decreaseswith a decrease in the opening area of the intake valve, and theinternal combustion engine is controlled based on the estimated degreeof turbulence in the cylinder. More specifically, the degree ofturbulence in the cylinder increases with a decrease in the opening areaof the intake valve, the degree of turbulence in the cylinder beingestimated based on the opening area of the intake valve, and theinternal combustion engine is so controlled that the ignition timing isdelayed with an increase in the degree of turbulence in the cylinder.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view schematically illustrating a first embodiment of adevice for controlling an internal combustion engine according to thepresent invention;

FIG. 2 is a view illustrating, in detail, an intake system of the devicefor controlling an internal combustion engine shown in FIG. 1;

FIG. 3 is a view illustrating in detail a cam for an intake valve and acam shaft shown in FIG. 1;

FIG. 4 is a view illustrating in detail a device for changing thevalve-lifting amount shown in FIG. 1;

FIG. 5 is a diagram illustrating a change in the valve-lifting amount ofthe intake valve accompanying the operation of the device for changingthe valve-lifting amount;

FIG. 6 is a view illustrating in detail an opening/closing timingshifting device shown in FIG. 1;

FIG. 7 is a diagram illustrating how the opening/closing timing of theintake valve shifts accompanying the operation of the opening/closingtiming shifting device;

FIG. 8 is a flowchart illustrating a method of calculating the ignitiontiming according to the first embodiment;

FIG. 9 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the valve-lifting amount LT and the pressure PM in the intakepipe;

FIG. 10 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the working angle VA and the pressure PM in the intake pipe;

FIG. 11 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the working angle VA and the pressure PM in the intake pipe;

FIG. 12 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the opening/closing timing (phase) VT and the pressure PM in theintake pipe;

FIG. 13 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the opening/closing timing (phase) VT and the pressure PM in theintake pipe;

FIG. 14 is a diagram illustrating a relationship between the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter and the engine rotational speed NE;

FIG. 15 is a diagram illustrating a relationship among the ignitiontiming SA, the pressure PCYL in the cylinder at the compression top deadcenter, and the intake air amount GN taken in by the cylinder per onerevolution;

FIG. 16 is a diagram illustrating a relationship between the ignitiontiming SA and the engine rotational speed NE;

FIG. 17 is a flowchart illustrating a method of calculating the amountof fuel injection according to a second embodiment;

FIG. 18 is a diagram illustrating a relationship among the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter, the valve-lifting amount LT and the pressure PM in the intakepipe;

FIG. 19 is a diagram illustrating a relationship among the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter, the working angle VA and the pressure PM in the intake pipe;

FIG. 20 is a diagram illustrating a relationship among the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter, the opening/closing timing (phase) VT and the pressure PM in theintake pipe;

FIG. 21 is a diagram illustrating a relationship between the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter and the engine rotational speed NE;

FIG. 22 is a diagram illustrating a relationship among the fuelinjection amount QINJ, the pressure PCYLIN in the cylinder at the intakebottom dead center, and the opening/closing timing (phase, valveoverlapping) VT;

FIG. 23 is a flowchart illustrating a method of calculating the ignitiontiming according to a third embodiment;

FIG. 24 is a diagram illustrating a relationship among the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center, the valve-lifting amount LT and the opening/closingtiming (phase) VT;

FIG. 25 is a diagram illustrating a relationship among the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center, the valve-lifting amount LT and the opening/closingtiming (phase) VT;

FIG. 26 is a diagram illustrating a relationship among the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center, the valve-lifting amount LT and the working angle VA;

FIG. 27 is a diagram illustrating a relationship among the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center, the valve-lifting amount LT and the working angle VA;

FIG. 28 is a diagram illustrating a relationship between the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center and the pressure PM in the intake pipe;

FIG. 29 is a diagram illustrating a relationship between the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center and the engine rotational speed NE;

FIG. 30 is a diagram illustrating a relationship among the correctedheat value KTWALL, the difference between the cylinder wall temperatureTwall and the normal condition of temperature TCYLb of gas in thecylinder at the compression top dead center, and the engine rotationalspeed NE;

FIG. 31 is a diagram illustrating a relationship among the intake airtemperature correction value KTIN, the engine cooling water temperatureT_(w), and the intake air amount Ga;

FIG. 32 is a diagram illustrating a relationship between the internalEGR gas temperature correction value KTEGR and the ratio of the internalEGR gas;

FIG. 33 is a diagram illustrating a relationship among the internal EGRgas temperature correction value KTEGR, the ignition timing of theprevious time and the amount of gas burnt per one revolution of theprevious time;

FIG. 34 is a diagram illustrating a relationship between the internalEGR gas temperature correction value KTEGR and the air-fuel ratio of theprevious time;

FIG. 35 is a diagram illustrating a relationship among the ignitiontiming SA, the temperature TCYL of gas in the cylinder at thecompression top dead center and the intake air amount GN per arevolution;

FIG. 36 is a flowchart illustrating a method of calculating the ignitiontiming according to a fourth embodiment;

FIG. 37 is a diagram illustrating a relationship among the normalcondition steady-state ratio KEGRb of the internal EGR gas, thevalve-lifting amount LT and the opening/closing timing (phase) VT;

FIG. 38 is a diagram illustrating a relationship among the normalcondition steady-state ratio KEGRb of the internal EGR gas ratio, theworking angle VA and the opening/closing timing (phase) VT;

FIG. 39 is a diagram illustrating a relationship between the normalcondition steady-state ratio KEGRb of the internal EGR gas ratio and thepressure PM in the intake pipe;

FIG. 40 is a diagram illustrating a relationship between the normalcondition steady-state ratio KEGRb of the internal EGR gas ratio and theengine rotational speed NE;

FIG. 41 is a diagram illustrating a relationship between the atmosphericpressure correction coefficient KPA and the atmospheric pressure;

FIG. 42 is a diagram illustrating a relationship among the backpressure, the engine rotational speed NE and the intake air amount GNper one revolution;

FIG. 43 is a diagram illustrating a relationship between the backpressure and the back pressure correction coefficient for correcting theinternal EGR gas ratio;

FIG. 44 is a diagram illustrating a relationship among the amount of theblown-back gas, the average opening area of the intake valve 2 (averageopening area of the intake valve during the valve overlapping period)and the average pressure differential before and after the intake valve2 (average differential between the pressure in the cylinder and thepressure in the intake pipe during the valve overlapping period);

FIG. 45 is a diagram illustrating a relationship between the amount ofthe blown-back gas and the steady-state ratio KEGRST of the internal EGRgas;

FIG. 46 is a diagram illustrating a relationship among the degree ofeffect due to the internal EGR gas of the previous time (=1-ratio ofchange KEGRSM from the previous time), the ratio KEGRO of the internalEGR gas in the previous time and the pressure PM in the intake pipe;

FIG. 47 is a diagram illustrating a relationship among the ignitiontiming SA, the ratio KEGR of the internal EGR gas and the intake airamount GN per a revolution;

FIG. 48 is a diagram illustrating a relationship between the ignitiontiming SA and the engine rotational speed NE;

FIG. 49 is a flowchart illustrating a method of calculating the ignitiontiming according to a fifth embodiment;

FIG. 50 is a diagram illustrating a relationship among the turbulenceCYLTRB in the cylinder, the valve-lifting amount LT and theopening/closing timing (phase) VT;

FIG. 51 is a diagram illustrating a relationship among the turbulenceCYLTRB in the cylinder, the working angle VA and the opening/closingtiming (phase) VT;

FIG. 52 is a diagram illustrating a relationship between the turbulenceCYLTRB in the cylinder and the pressure PM in the intake pipe;

FIG. 53 is a diagram illustrating a relationship between the turbulenceCYLTRB in the cylinder and the engine rotational speed NE;

FIG. 54 is a diagram illustrating a relationship among the ignitiontiming SA, the turbulence CYLTRB in the cylinder and the air intakeamount GN per a revolution;

FIG. 55 is a diagram illustrating a relationship between the ignitiontiming SA and the engine rotational speed NE;

FIG. 56 is a flowchart illustrating a method of controlling the camaccording to a sixth embodiment;

FIG. 57 is a diagram illustrating a relationship among the acceleratoropening degree, the engine rotational speed and the cam to be selected;

FIG. 58 is a diagram illustrating a relationship among the delay inchanging the cam, the engine rotational speed and the cooling watertemperature;

FIG. 59 is a diagram illustrating a relationship between the delay inchanging the cam and the hydraulic pressure;

FIG. 60 is a diagram illustrating a relationship between the timing forproducing an instruction for changing the cam and the timing at whichthe cam really changes;

FIG. 61 is a flowchart illustrating a method of calculating the amountof fuel injection according to a sixth embodiment;

FIG. 62 is a diagram illustrating a relationship among the responsecorrection coefficient, the type of cam, the engine rotational speed andthe intake air amount GN per one revolution;

FIG. 63 is a diagram illustrating a relationship between the amount offuel injection and the intake air amount per one revolution;

FIG. 64 is a flowchart illustrating a routine for calculating theignition timing according to the sixth embodiment; and

FIG. 65 is a diagram illustrating a relationship among the ignitiontiming, the type of cam, the engine rotational speed and the intake airamount GN per one revolution.

BEST MODE FOR CARRYING OUT THE INVENTION

Embodiments of the invention will now be described with reference to theaccompanying drawings.

FIG. 1 is a view schematically illustrating a first embodiment of adevice for controlling an internal combustion engine according to thepresent invention, and FIG. 2 is a view illustrating, in detail, anintake system of the device for controlling an internal combustionengine shown in FIG. 1. In FIGS. 1 and 2, reference numeral 1 denotes aninternal combustion engine, 2 denotes intake valves, 3 denotes exhaustvalves, 4 denotes cams for opening and closing the intake valves, 5denotes cams for opening and closing the exhaust valves, 6 denotes a camshaft supporting the cams 4 for intake valves, and 7 denotes a cam shaftsupporting the cams 5 for exhaust valves. FIG. 3 is a view illustratingin detail the cam for the intake valve and the cam shaft shown inFIG. 1. As shown in FIG. 3, the cam 4 according to this embodiment has acam profile that is changing in the direction of the center axis of thecam shaft. That is, the cam 4 according to this embodiment has a nose atthe left end in FIG. 3 which is higher than a nose at the right end.That is, the valve-lifting amount of the intake valve 2 according tothis embodiment is smaller when the valve lifter is in contact with theright end of the cam 4 than when the valve lifter is in contact with theleft end of the cam 4.

Reverting to FIGS. 1 and 2, reference numeral 8 denotes a combustionchamber formed in the cylinder, and 9 denotes a device for changing thevalve-lifting amount by moving the cam 4 in a direction of the centeraxis of the cam shaft in order to change the valve-lifting amount.Namely, upon operating the device 9 for changing the valve-liftingamount, the valve lifter is brought into contact with the cam 4 at theleft end (FIG. 3) of the cam 4 or the valve lifter is brought intocontact with the cam 4 at the right end (FIG. 3) of the cam 4. When thevalve-lifting amount of the intake valve 2 is changed by the device 9for changing the valve-lifting amount, the opening area of the intakevalve 2 changes. With the intake valve 2 of this embodiment, the openingarea of the intake valve 2 increases with an increase in thevalve-lifting amount. Reference numeral 10 denotes a driver for drivingthe device 9 for changing the valve-lifting amount, and 11 denotes anopening/closing timing shifting device for shifting the opening/closingtiming of the intake valve without changing the valve-opening period ofthe intake valve 2. Namely, by operating the opening/closing timingshifting device 11, the opening/closing timing of the intake valve 2 canbe shifted toward the advancing side or toward the delaying side.Reference numeral 12 denotes an oil control valve for controlling thehydraulic pressure for actuating the opening/closing timing shiftingdevice 11. The variable valve mechanism according to this embodimentincludes both the device 9 for changing the valve-lifting amount and theopening/closing timing shifting device 11.

Reference numeral 13 denotes a crank shaft, 14 denotes an oil pan, 15denotes a fuel injection valve, 16 denotes a sensor for detecting thevalve-lifting amount of the intake valve 2 and the amount of shiftingthe opening/closing timing, and reference numeral 17 denotes a sensorfor detecting the engine rotational speed. Reference numeral 18 denotesan intake pipe pressure sensor for detecting the pressure in the intakepipe through which the intake air is fed into the cylinder, 19 denotesan air flow meter, 20 denotes a cooling water temperature sensor fordetecting the temperature of the internal combustion engine coolingwater, 21 denotes an intake air temperature sensor for detecting thetemperature of the intake air fed into the cylinder through the intakepipe, and 22 denotes an ECU (electronic control unit). Reference numeral50 denotes a cylinder, 51 and 52 denote intake pipes, 53 denotes a surgetank, 54 denotes an exhaust pipe, and 55 denotes a spark plug.

FIG. 4 is a view illustrating in detail the device for changing thevalve-lifting amount shown in FIG. 1. In FIG. 4, reference numeral 30denotes a magnetic material coupled to the cam shaft 6 for the intakevalves, 31 denotes a coil for urging the magnetic material 30 toward theleft, and 32 denotes a compression spring for urging the magneticmaterial 30 toward the right. As the amount of electric current suppliedto the coil 31 increases, the cam 4 and the cam shaft 6 move toward theleft by an increased amount, and the valve-lifting amount of the intakevalve 2 decreases.

FIG. 5 is a diagram illustrating a change in the valve-lifting amount ofthe intake valve accompanying the operation of the device for changingthe valve-lifting amount. Referring to FIG. 5, as the amount of electriccurrent supplied to the coil 31 decreases, the valve-lifting amount ofthe intake valve 2 increases (solid line→broken line→dot-dash chainline) In this embodiment, further, the valve-opening period of theintake valve 2 varies accompanying the operation of the device 9 forchanging the valve-lifting amount. Namely, the working angle of theintake valve 2 changes, too. If described in detail, the working angleof the intake valve 2 increases (solid line→broken line→dot-dash chainline) accompanying an increase in the valve-lifting amount of the intakevalve 2. In this embodiment, further, the timing at which thevalve-lifting amount of the intake valve 2 becomes a peak also variesaccompanying the operation of the device 9 for changing thevalve-lifting amount. If described in detail, the timing at which thevalve-lifting amount of the intake valve 2 becomes a peak is delayed(solid line→broken line→dot-dash chain line) accompanying an increase inthe valve-lifting amount of the intake valve 2.

FIG. 6 is a view illustrating in detail the opening/closing timingshifting device shown in FIG. 1. In FIG. 6, reference numeral 40 denotesa fluid passage on the advancing side for shifting the opening/closingtiming of the intake valve 2 toward the advancing side, referencenumeral 41 denotes a fluid passage on the delaying side for shifting theopening/closing timing of the intake valve 2 toward the delaying side,and 42 denotes an oil pump. As the hydraulic pressure increases in thefluid passage 40 on the advancing side, the opening/closing timing ofthe intake valve 2 is shifted toward the advancing side. Namely, therotational phase of the cam shaft 6 is advanced with respect to thecrank shaft 13. As the hydraulic pressure increases in the fluid passage41 on the delaying side, on the other hand, the opening/closing timingof the intake valve 2 is shifted toward the delaying side. Namely, therotational phase of the cam shaft 6 is delayed with respect to the crankshaft 13.

FIG. 7 is a diagram illustrating how the opening/closing timing of theintake valve shifts accompanying the operation of the opening/closingtiming shifting device. As the hydraulic pressure increases in the fluidpassage 40 on the advancing side as shown in FIG. 7, the opening/closingtiming of the intake valve 2 is shifted toward the advancing side (solidline→broken line→dot-dash chain line). Here, the valve-opening period ofthe intake valve 2 remains unchanged. Namely, there is no change in thelength of period in which the intake valve 2 remains opened.

As the valve-lifting amount of the intake valve 2, working angle and theopening/closing timing (phase) are varied by the device 9 for changingthe valve-lifting amount and by the opening/closing timing shiftingdevice 11 as described above, then, the pressure in the cylinder varies.If the ignition is conducted at a predetermined timing irrespective of achange in the pressure in the cylinder, an optimum ignition timing isnot accomplished, and the internal combustion engine is not suitablycontrolled. In order to conduct the ignition at an optimum timing and tosuitably control the internal combustion engine, therefore, the pressurein the cylinder must be correctly calculated depending upon changes inthe valve-lifting amount of the intake valve 2, upon the working angleand upon the opening/closing timing (phase) thereof.

FIG. 8 is a flowchart illustrating a method of calculating the ignitiontiming according to the embodiment. This routine is executed atpredetermined time intervals. When the routine starts as shown in FIG.8, it is, first, judged at step 100 if the engine is being started. Whenthe result is YES, the pressure in the cylinder is correctly calculatedat the start of the engine where the amount of the fuel is beingincreased, it is so judged based thereupon that there is no need todetermine the ignition timing, and the routine ends. When the result isNO, on the other hand, the routine proceeds to step 101. At step 101,the normal condition of pressure PCYLb in the cylinder at thecompression top dead center is calculated based on the valve-liftingamount LT of the intake valve 2, working angle VA, opening/closingtiming VT, pressure PM in the intake pipe and the engine rotationalspeed NE.

FIG. 9 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the valve-lifting amount LT and the pressure PM in the intakepipe. As shown in FIG. 9, the normal condition of pressure PCYLb in thecylinder at the compression top dead center calculated at step 101increases with an increase in the valve-lifting amount LT, or increaseswith an increase in the pressure PM in the intake pipe. FIG. 10 is adiagram illustrating a relationship among the normal condition ofpressure PCYLb in the cylinder at the compression top dead center, theworking angle VA and the pressure PM in the intake pipe. As shown inFIG. 10, the normal condition of pressure PCYLb in the cylinder at thecompression top dead center calculated at step 101 increases with adecrease in the working angle VA when the intake valve 2 is fully closedafter the intake bottom dead center. FIG. 11 is a diagram illustrating arelationship among the normal condition of pressure PCYLb in thecylinder at the compression top dead center, the working angle VA andthe pressure PM in the intake pipe. As shown in FIG. 11, the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter calculated at step 101 increases with an increase in the workingangle VA when the intake valve 2 is fully closed before the intakebottom dead center.

FIG. 12 is a diagram illustrating a relationship among the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter, the opening/closing timing (phase) VT and the pressure PM in theintake pipe. As shown in FIG. 12, the normal condition of pressure PCYLbin the cylinder at the compression top dead center calculated at step101 increases as the opening/closing timing (phase) VT advances when theintake valve 2 is fully closed after the intake bottom dead center. FIG.13 is a diagram illustrating a relationship among the normal conditionof pressure PCYLb in the cylinder at the compression top dead center,the opening/closing timing (phase) VT and the pressure PM in the intakepipe. As shown in FIG. 13, the normal condition of pressure PCYLb in thecylinder at the compression top dead center calculated at step 101increases as the opening/closing timing (phase) VT delays when theintake valve 2 is fully closed before the intake bottom dead center.FIG. 14 is a diagram illustrating a relationship between the normalcondition of pressure PCYLb in the cylinder at the compression top deadcenter and the engine rotational speed NE. As shown in FIG. 14, thenormal condition of pressure PCYLb in the cylinder at the compressiontop dead center calculated at step 101 becomes a peak when the enginerotational speed NE is an intermediate speed.

Reverting to the description of FIG. 8, a pressure PCYL in the cylinderat the compression top dead center is calculated at step 102 from thenormal condition of pressure PCYLb in the cylinder at the compressiontop dead center under the present engine operating conditions. Then, atstep 103, an ignition timing SA is calculated based on the pressure PCYLin the cylinder at the compression top dead center, engine rotationalspeed NE, and intake air amount GN taken in by the cylinder per arevolution, i.e., the intake air amount GN taken in by the cylinderduring one time of intake stroke. FIG. 15 is a diagram illustrating arelationship among the ignition timing SA, the pressure PCYL in thecylinder at the compression top dead center, and the intake air amountGN taken in by the cylinder per one revolution. As shown in FIG. 15, theignition timing SA calculated at step 103 is delayed as the pressurePCYL increases in the cylinder at the compression top dead center, andis delayed with an increase in the intake air amount GN taken in by thecylinder per one revolution. FIG. 16 is a diagram illustrating arelationship between the ignition timing SA and the engine rotationalspeed NE. As shown in FIG. 16, the ignition timing SA calculated at step103 advances with an increase in the engine rotational speed NE.

In this embodiment as described above, a pressure in the cylinder(pressure PCYL in the cylinder at the compression top dead center) iscalculated at steps 101 and 102 in FIG. 8 based on the opening area ofthe intake valve 2 that varies depending upon the valve-lifting amountLT varied by the device 9 for changing the valve-lifting amount which isthe variable valve mechanism, and the internal combustion engine iscontrolled based on the pressure in the cylinder. According to thisembodiment, therefore, the internal combustion engine can be controlledbased not only upon the peak combustion pressure in the cylinder butalso upon the pressure in the cylinder at a moment other than the peakcombustion pressure unlike the case of detecting the pressure in thecylinder by using the cylinder pressure sensor employed by the devicefor controlling internal combustion engines taught in JapaneseUnexamined Patent Publication (Kokai) No. 9-53503. The internalcombustion engine can be suitably controlled even when the opening areaof the intake valve 2 is varied. More specifically, the pressure in thecylinder calculated based on the opening area of the intake valveincreases with an increase in the opening area of the intake valve 2 asshown in FIG. 9, and the internal combustion engine is so controlledthat the ignition timing SA is delayed with an increase in the pressurein the cylinder as shown in FIG. 15.

In this embodiment, further, the pressure in the cylinder (pressure PCYLin the cylinder at the compression top dead center) is calculated atsteps 101 and 102 in FIG. 8 based on the working angle VA of the intakevalve 2 that is varied by the device 9 for changing the valve-liftingamount which is the variable valve mechanism, and the internalcombustion engine is controlled based on the pressure in the cylinder.According to this embodiment, therefore, the internal combustion enginecan be controlled based not only upon the peak combustion pressure inthe cylinder but also upon the pressure in the cylinder at a momentother than the peak combustion pressure unlike the case of detecting thepressure in the cylinder by using the cylinder pressure sensor employedby the device for controlling internal combustion engines taught inJapanese Unexamined Patent Publication (Kokai) No. 9-53503. The internalcombustion engine can be suitably controlled even when the working angleVA of the intake valve 2 is varied. More specifically, the pressure inthe cylinder calculated based on the working angle VA of the intakevalve 2 increases with a decrease in the working angle VA of the intakevalve 2 when the intake valve 2 is fully closed after the intake bottomdead center as shown in FIG. 10, and the internal combustion engine isso controlled that the ignition timing SA is delayed with an increase inthe pressure in the cylinder as shown in FIG. 15. Further, the pressurein the cylinder calculated based on the working angle VA of the intakevalve 2 increases with an increase in the working angle VA of the intakevalve 2 when the intake valve 2 is fully closed before the intake bottomdead center as shown in FIG. 11, and the internal combustion engine isso controlled that the ignition timing SA is delayed with an increase inthe pressure in the cylinder as shown in FIG. 15.

In this embodiment, further, the pressure in the cylinder (pressure PCYLin the cylinder at the compression top dead center) is calculated atsteps 101 and 102 in FIG. 8 based on both the opening area and theworking angle VA of the intake valve 2 that are varied by the device 9for changing the valve-lifting amount which is the variable valvemechanism, and the internal combustion engine is controlled based on thepressure in the cylinder. According to this embodiment, therefore, theinternal combustion engine can be suitably controlled by more correctlycalculating the pressure in the cylinder than when the pressure in thecylinder is calculated based only upon the opening area of the intakevalve 2 but not upon the working angle VA of the intake valve 2, or thanwhen the pressure in the cylinder is calculated based only upon theworking angle VA of the intake valve 2 but not upon the opening area ofthe intake valve 2.

In this embodiment, further, the pressure in the cylinder (pressure PCYLin the cylinder at the compression top dead center) is calculated atsteps 101 and 102 in FIG. 8 based upon the opening/closing timing(phase) VT of the intake valve 2, pressure PM in the intake pipe andengine rotational speed NE, and the internal combustion engine iscontrolled based on the pressure in the cylinder. According to thisembodiment, therefore, the internal combustion engine can be suitablycontrolled by more correctly calculating the pressure in the cylinderthan when the pressure in the cylinder is not calculated based on theopening/closing timing (phase) VT of the intake valve 2, pressure PM inthe intake pipe and engine rotational speed NE.

In this embodiment, the pressure in the cylinder is calculated based onthe opening area of the intake valve and the like, and the internalcombustion engine is controlled based on the pressure in the cylinder.According to another embodiment, the pressure in the cylinder iscalculated based on the opening areas of the exhaust valves, and theinternal combustion engine is controlled based on the pressure in thecylinder. Namely, the invention can be applied not only to the intakevalves but also to the exhaust valves.

Described below is a second embodiment of the device for controlling aninternal combustion engine according to the invention. The constitutionof this embodiment is nearly the same as the constitution of the firstembodiment illustrated in FIGS. 1 to 7. In this embodiment, too, thepressure in the cylinder varies as the valve-lifting amount of theintake valve 2, working angle and opening/closing timing (phase) arevaried by the device 9 for changing the valve-lifting amount and by theopening/closing timing shifting device 11. If the amount of fuelinjection is set to be constant irrespective of a change in the pressurein the cylinder, the real air-fuel ratio deviates from a target air-fuelratio, and the internal combustion engine is not suitably controlled. Inorder to calculate an optimum amount of fuel injection and to suitablycontrol the internal combustion engine, therefore, the pressure in thecylinder must be correctly calculated depending upon the valve-liftingamount of the intake valve 2, upon the working angle and upon theopening/closing timing (phase) thereof.

FIG. 17 is a flowchart illustrating a method of calculating the amountof fuel injection according to the embodiment. This routine is executedat predetermined time intervals. When the routine starts as shown inFIG. 17, it is first judged at step 200 if the engine is being started.When the result is YES, the amount of fuel injection is determinedirrespective of the pressure in the cylinder at the start of the enginewhere the amount of the fuel is being increased, it is so judged basedthereupon that there is no need to correctly calculate the pressure inthe cylinder for determining the amount of fuel injection, and theroutine ends. When the result is NO, on the other hand, the routineproceeds to step 201. At step 201, the normal condition of pressurePCYLINb in the cylinder at the intake bottom dead center is calculatedbased on the valve-lifting amount LT of the intake valve 2, workingangle VA, opening/closing timing VT, pressure PM in the intake pipe andthe engine rotational speed NE.

FIG. 18 is a diagram illustrating a relationship among the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter, the valve-lifting amount LT and the pressure PM in the intakepipe. As shown in FIG. 18, the normal condition of pressure PCYLINb inthe cylinder at the intake bottom dead center calculated at step 201increases with an increase in the valve-lifting amount LT, or increaseswith an increase in the pressure PM in the intake pipe. FIG. 19 is adiagram illustrating a relationship among the normal condition ofpressure PCYLINb in the cylinder at the intake bottom dead center, theworking angle VA and the pressure PM in the intake pipe. As shown inFIG. 19, the normal condition of pressure PCYLINb in the cylinder at theintake bottom dead center calculated at step 201 increases with adecrease in the working angle VA.

FIG. 20 is a diagram illustrating a relationship among the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter, the opening/closing timing (phase) VT and the pressure PM in theintake pipe. As shown in FIG. 20, the normal condition of pressurePCYLINb in the cylinder at the intake bottom dead center calculated atstep 201 increases as the opening/closing timing (phase) VT advances.FIG. 21 is a diagram illustrating a relationship between the normalcondition of pressure PCYLINb in the cylinder at the intake bottom deadcenter and the engine rotational speed NE. As shown in FIG. 21, thenormal condition of pressure PCYLINb in the cylinder at the intakebottom dead center calculated at step 201 becomes a peak when the enginerotational speed NE is an intermediate speed.

Reverting to the description of FIG. 17, the pressure PCYLIN in thecylinder at the intake bottom dead center is calculated at step 202 fromthe normal condition of pressure PCYLINb in the cylinder at the intakebottom dead center under the present engine operating conditions. Then,at step 203, the fuel injection amount QINJ is calculated based on thepressure PCYLIN in the cylinder at the intake bottom dead center andopening/closing timing (phase, valve overlapping) VT. FIG. 22 is adiagram illustrating a relationship among the fuel injection amountQINJ, the pressure PCYLIN in the cylinder at the intake bottom deadcenter, and the opening/closing timing (phase, vave overlapping) VT. Asshown in FIG. 22, the fuel injection amount QINJ calculated at step 203increases as the pressure PCYLIN increases in the cylinder at the intakebottom dead center, and increases as the opening/closing timing (phase)VT is delayed, i.e., as the valve overlapping period between the intakevalve 2 and the exhaust valve 3 decreases.

In this embodiment as described above, the pressure in the cylinder(pressure PCYLIN in the cylinder at the intake bottom dead center) iscalculated at steps 201 and 202 in FIG. 17 based on the opening area ofthe intake valve 2 that varies depending upon the valve-lifting amountLT varied by the device 9 for changing the valve-lifting amount which isthe variable valve mechanism, and the internal combustion engine iscontrolled based on the pressure in the cylinder. According to thisembodiment, therefore, the internal combustion engine can be controlledbased not only upon the peak combustion pressure in the cylinder butalso upon the pressure in the cylinder at a moment other than the peakcombustion pressure unlike the case of detecting the pressure in thecylinder by using the cylinder pressure sensor employed by the devicefor controlling internal combustion engines taught in JapaneseUnexamined Patent Publication (Kokai) No. 9-53503. The internalcombustion engine can be suitably controlled even when the opening areaof the intake valve 2 is varied. More specifically, the pressure in thecylinder calculated based on the opening area of the intake valveincreases with an increase in the opening area of the intake valve 2 asshown in FIG. 18, and the internal combustion engine is so controlledthat the fuel injection amount QINJ increases with an increase in thepressure in the cylinder as shown in FIG. 22.

In this embodiment, further, the pressure in the cylinder (pressurePCYLIN in the cylinder at the intake bottom dead center) is calculatedat steps 201 and 202 in FIG. 17 based on the working angle VA of theintake valve 2 that is varied by the device 9 for changing thevalve-lifting amount which is the variable valve mechanism, and theinternal combustion engine is controlled based on the pressure in thecylinder. According to this embodiment, therefore, the internalcombustion engine can be controlled based not only upon the peakcombustion pressure in the cylinder but also upon the pressure in thecylinder at a moment other than the peak combustion pressure unlike thecase of detecting the pressure in the cylinder by using the cylinderpressure sensor employed by the device for controlling internalcombustion engines taught in Japanese Unexamined Patent Publication(Kokai) No. 9-53503. The internal combustion engine can be suitablycontrolled even when the working angle VA of the intake valve 2 isvaried. More specifically, the pressure in the cylinder calculated basedon the working angle VA of the intake valve 2 increases with a decreasein the working angle VA of the intake valve 2 as shown in FIG. 19, andthe internal combustion engine is so controlled that the fuel injectionamount QINJ increases with an increase in the pressure in the cylinderas shown in FIG. 22.

In this embodiment, further, the pressure in the cylinder (pressurePCYLIN in the cylinder at the intake bottom dead center) is calculatedat steps 201 and 202 in FIG. 17 based on both the opening area and theworking angle VA of the intake valve 2 that are varied by the device 9for changing the valve-lifting amount which is the variable valvemechanism, and the internal combustion engine is controlled based on thepressure in the cylinder. According to this embodiment, therefore, theinternal combustion engine can be suitably controlled by more correctlycalculating the pressure in the cylinder than when the pressure in thecylinder is calculated based only upon the opening area of the intakevalve 2 but not upon the working angle VA of the intake valve 2, or thanwhen the pressure in the cylinder is calculated based only upon theworking angle VA of the intake valve 2 but not upon the opening area ofthe intake valve 2.

In this embodiment, further, the pressure in the cylinder (pressurePCYLIN in the cylinder at the intake bottom dead center) is calculatedat steps 201 and 202 in FIG. 17 based on the opening/closing timing(phase) VT of the intake valve 2, pressure PM in the intake pipe andengine rotational speed NE, and the internal combustion engine iscontrolled based on the pressure in the cylinder. According to thisembodiment, therefore, the internal combustion engine can be suitablycontrolled by more correctly calculating the pressure in the cylinderthan when the pressure in the cylinder is not calculated based on theopening/closing timing (phase) VT of the intake valve 2, pressure PM inthe intake pipe and engine rotational speed NE.

In this embodiment, the pressure in the cylinder is calculated based onthe opening area of the intake valve and the like, and the internalcombustion engine is controlled based on the pressure in the cylinder.According to another embodiment, the pressure in the cylinder iscalculated based on the opening areas of the exhaust valves, and theinternal combustion engine is controlled based on the pressure in thecylinder. Namely, the invention can be applied not only to the intakevalves but also to the exhaust valves.

Described below is a third embodiment of the device for controlling aninternal combustion engine according to the invention. The constitutionof this embodiment is nearly the same as the constitution of the firstembodiment illustrated in FIGS. 1 to 7. The temperature of gas in thecylinder varies as the valve-lifting amount of the intake valve 2,working angle and opening/closing timing (phase) are varied by thedevice 9 for changing the valve-lifting amount and by theopening/closing timing shifting device 11. If the ignition is conductedat a predetermined timing irrespective of a change in the temperature ofgas in the cylinder, an optimum ignition timing is not accomplished, andthe internal combustion engine is not suitably controlled. In order toconduct the ignition at an optimum timing and to suitably control theinternal combustion engine, therefore, the temperature of gas in thecylinder must be correctly calculated depending upon changes in thevalve-lifting amount of the intake valve 2, upon the working angle andupon the opening/closing timing (phase) thereof.

FIG. 23 is a flowchart illustrating a method of calculating the ignitiontiming according to the embodiment. This routine is executed atpredetermined time intervals. When the routine starts as shown in FIG.23, it is first judged at step 300 if the engine is being started. Whenthe result is YES, the temperature of gas in the cylinder is correctlycalculated at the start of the engine where the amount of the fuel isbeing increased, it is so judged based thereupon that there is no needof determining the ignition timing, and the routine ends. When theresult is NO, on the other hand, the routine proceeds to step 301. Atstep 301, the normal condition of temperature TCYLb of gas in thecylinder at the compression top dead center is calculated based on thevalve-lifting amount LT of the intake valve 2, working angle VA,opening/closing timing VT, pressure PM in the intake pipe and the enginerotational speed NE.

FIG. 24 is a diagram illustrating a relationship among the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center, the valve-lifting amount LT and the opening/closingtiming (phase) VT. As shown in FIG. 24, the normal condition oftemperature TCYLb of gas in the cylinder at the compression top deadcenter calculated at step 301 increases with an increase in thevalve-lifting amount LT, or increases as the opening/closing timing(phase) VT advances when the intake valve 2 is fully closed after theintake bottom dead center. As shown in FIG. 25, the normal condition oftemperature TCYLb of gas in the cylinder at the compression top deadcenter calculated at step 301 increases with an increase in thevalve-lifting amount LT, or increases as the opening/closing timing(phase) VT is delayed when the intake valve 2 is fully closed before theintake bottom dead center. FIG. 26 is a diagram illustrating arelationship among the normal condition of temperature TCYLb of gas inthe cylinder at the compression top dead center, the valve-liftingamount LT and the working angle VA. As shown in FIG. 26, the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center calculated at step 301 increases with an increase in theworking angle VA when the intake valve 2 is fully closed after theintake bottom dead center. FIG. 27 is a diagram illustrating arelationship among the normal condition of temperature TCYLb of gas inthe cylinder at the compression top dead center, the valve-liftingamount LT and the working angle VA. As shown in FIG. 27, the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center calculated at step 301 increases with a decrease in theworking angle VA when the intake valve 2 is fully closed before theintake bottom dead center.

FIG. 28 is a diagram illustrating a relationship between the normalcondition of temperature TCYLb of gas in the cylinder at the compressiontop dead center and the pressure PM in the intake pipe. As shown in FIG.28, the normal condition of temperature TCYLb of gas in the cylinder atthe compression top dead center calculated at step 301 increases with anincrease in the pressure PM in the intake pipe. As shown in FIG. 29, thenormal condition of temperature TCYLb of gas in the cylinder at thecompression top dead center calculated at step 301 becomes a peak whenthe engine rotational speed NE is an intermediate speed.

Reverting to the description of FIG. 23, a corrected heat value KTWALLis calculated at step 302 based on the cylinder wall temperature Twall.The cylinder wall temperature Twall is estimated in compliance with thefollowing formula,Twall=(K 1×Ga(i)−Tw(i)−Twall(i−1))×K 2+Twall(i)

where K1 is a combustion correction coefficient, K2 is a

response coefficient, Ga is an intake air amount calculatedbased on theoutput of the air flow meter 19, Tw is the temperature of the enginecooling water, i is a value of

when the routine shown in FIG. 23 is executed this time,

and i-i is a value of when the routine shown in FIG. 23 was executed inthe previous time.

The combustion correction coefficient K1 assumes a positive value whenthe fuel injected from the fuel injection valve 15 is burning to therebycut the fuel, and assumes a negative value during the motoring in whichno fuel is burning. FIG. 30 is a diagram illustrating a relationshipamong the corrected heat value KTWALL, the difference between thecylinder wall temperature Twall and the normal condition of temperatureTCYLb of gas in the cylinder at the compression top dead center, and theengine rotational speed NE. As shown in FIG. 30, the corrected heatvalue KTWALL increases as the cylinder wall temperature Twall becomeshigher than the normal condition of temperature TCYLb of gas in thecylinder at the compression top dead center, and increases as the enginerotational speed NE decreases.

Reverting to the description of FIG. 23, an intake air temperaturecorrection value KTIN is calculated at step 303 based on the temperatureof the intake air taken in by the cylinder.

FIG. 31 is a diagram illustrating a relationship among the intake airtemperature correction value KTIN, the engine cooling water temperatureTw and the intake air amount Ga. As shown in FIG. 31, the intake airtemperature correction value KTIN increases with an increase in theengine cooling water temperature Tw, and increases with a decrease inthe intake air amount Ga.

Reverting to the description of FIG. 23, an internal EGR gas temperaturecorrection value KTEGR is calculated at step 304 based on the ratio ofthe internal EGR gas in the cylinder. FIG. 32 is a diagram illustratinga relationship between the internal EGR gas temperature correction valueKTEGR and the ratio of the internal EGR gas. As shown in FIG. 32, theinternal EGR gas temperature correction value KTEGR increases with anincrease in the ratio of the internal EGR gas. As a modified example ofthis embodiment, it is possible to calculate the internal EGR gastemperature correction value KTEGR based on the amount of the internalEGR gas. In this case, the internal EGR gas temperature correction valueKTEGR increases with an increase in the amount of the internal EGR gas.As another modified example of this embodiment, it is allowable tocalculate the internal EGR gas temperature correction value KTEGR basedon the ignition timing of the previous time and the amount of the burntgas per a revolution of the previous time. FIG. 33 is a diagramillustrating a relationship among the internal EGR gas temperaturecorrection value KTEGR, the ignition timing of the previous time and theamount of burnt gas per a revolution of the previous time. As shown inFIG. 33, the internal EGR gas temperature correction value KTEGRincreases as the ignition timing of the previous time is delayed, andincreases with an increase in the amount of burnt gas per a revolutionof the previous time. As a further modified example of this embodiment,it is possible to calculate the internal EGR gas temperature correctionvalue KTEGR based on the air-fuel ratio of the previous time. FIG. 34 isa diagram illustrating a relationship between the internal EGR gastemperature correction value KTEGR and the air-fuel ratio of theprevious time. As shown in FIG. 34, the internal EGR gas temperaturecorrection value KTEGR becomes a peak at an air-fuel ratio which isslightly richer than the stoichiometric ratio, and decreases towardeither the rich side or the lean side.

Reverting to the description of FIG. 23, a temperature TCYL of gas inthe cylinder at the compression top dead center(TCYL←TCYLb×KTWALL×KTIN×KTEGR) is calculated at step 305 based upon thenormal condition of temperature TCYLb of gas in the cylinder at thecompression top dead center calculated at step 301, corrected heat valueKTWALL calculated at step 302, intake air temperature correction valueKTIN calculated at step 303 and internal EGR gas temperature correctionvalue KTEGR calculated at step 304. Then, at step 306, an ignitiontiming SA is calculated based upon the gas temperature TCYL at thecompression top dead center, intake air amount GN per one revolution andengine rotational speed NE. FIG. 35 is a diagram illustrating arelationship among the ignition timing SA, the temperature TCYL of gasin the cylinder at the compression top dead center and the intake airamount GN per a revolution. As shown in FIG. 35, the ignition timing SAcalculated at step 306 is delayed as the temperature TCYL of gas in thecylinder at the compression top dead center becomes high, and is delayedas the intake air amount GN increases per a revolution. As shown in FIG.16, further, the ignition timing SA calculated at step 306 advances asthe engine rotational speed NE increases.

In this embodiment as described above, the temperature of gas in thecylinder (temperature TCYL of gas in the cylinder at the compression topdead center) is calculated at steps 301 and 305 in FIG. 23 based uponthe opening area of the intake valve 2 that varies depending upon thevalve-lifting amount LT which is changed by the device 9 for changingthe valve-lifting amount, which is the variable valve mechanism, and theinternal combustion engine is controlled based upon the temperature ofgas in the cylinder. According to this embodiment, therefore, it ispossible to more suitably control the internal combustion engine thanwhen the internal combustion engine is controlled based upon thetemperature of the cylinder wall, that is done by the device forcontrolling internal combustion engines taught in Japanese UnexaminedPatent Publication (Kokai) No. 4-81574. The internal combustion enginecan be suitably controlled even when the opening area of the intakevalve 2 is varied. More specifically, the temperature of gas in thecylinder calculated based on the opening area of the intake valveincreases with an increase in the opening area of the intake valve 2 asshown in FIGS. 24 and 25, and the internal combustion engine is socontrolled that the ignition timing SA is delayed as the temperature ofgas in the cylinder increases as shown in FIG. 35.

In this embodiment, further, the temperature of gas in the cylinder(temperature TCYL of gas in the cylinder at the compression top deadcenter) is calculated at steps 301 and 305 in FIG. 23 based on theworking angle VA of the intake valve 2 that is varied by the device 9for changing the valve-lifting amount, which is the variable valvemechanism, and the internal combustion engine is controlled based on thetemperature of gas in the cylinder. According to this embodiment,therefore, it is possible to more suitably control the internalcombustion engine than when the internal combustion engine is controlledbased upon the temperature of the cylinder wall, which is done by thedevice for controlling internal combustion engines taught in JapaneseUnexamined Patent Publication (Kokai) No. 4-81574. The internalcombustion engine can be suitably controlled even when the working angleVA of the intake valve 2 is varied. More specifically, when the intakevalve 2 is fully closed after the suction bottom dead center as shown inFIG. 26, the temperature of gas in the cylinder calculated based on theworking angle VA of the intake valve 2 increases with an increase in theworking angle VA of the intake valve 2, and the internal combustionengine is so controlled that the ignition timing SA is delayed as thetemperature of gas in the cylinder increases as shown in FIG. 35.Further, when the intake valve 2 is fully closed before the suctionbottom dead center as shown in FIG. 27, the temperature of gas in thecylinder calculated based on the working angle VA of the intake valve 2increases with a decrease in the working angle VA of the intake valve 2,and the internal combustion engine is so controlled that the ignitiontiming SA is delayed as the temperature of gas in the cylinder increasesas shown in FIG. 35.

In this embodiment, further, the temperature of gas in the cylinder(temperature TCYL of gas in the cylinder at the compression top deadcenter) is calculated at steps 301 and 305 in FIG. 23 based on both theopening area and the working angle VA of the intake valve 2 that arevaried by the device 9 for changing the valve-lifting amount, which isthe variable valve mechanism, and the internal combustion engine iscontrolled based on the temperature of gas in the cylinder. According tothis embodiment, therefore, it is possible to more suitably control theinternal combustion engine by correctly calculating the temperature ofgas in the cylinder than when the temperature of gas in the cylinder iscalculated based only upon the opening area of the intake valve 2 butnot upon the working angle VA of the intake valve 2 or than when thetemperature of gas in the cylinder is calculated based only upon theworking angle VA of the intake valve 2 but not upon the opening area ofthe intake valve 2.

In this embodiment, further, the temperature of gas in the cylinder(temperature TCYL of gas in the cylinder at the compression top deadcenter) is calculated at steps 301 and 305 in FIG. 23 based upon theopening/closing timing (phase) VT of the intake valve 2, pressure PM inthe intake pipe and the engine rotational speed NE, and the internalcombustion engine is controlled based on the temperature of gas in thecylinder. According to this embodiment, therefore, it is possible tosuitably control the internal combustion engine by more correctlycalculating the temperature of gas in the cylinder than when thetemperature of gas in the cylinder is not calculated based theopening/closing timing (phase) VT of the intake valve 2, pressure PM inthe intake pipe, and engine rotational speed NE.

In this embodiment, further, the temperature of gas in the cylinder iscorrected at step 302 in FIG. 23 based on the cylinder wall temperatureTwall and the engine rotational speed NE, the temperature of gas in thecylinder is corrected at step 303 based on the intake air amount Ga, andthe temperature of gas in the cylinder is corrected at step 304 based onthe amount of the internal EGR gas (ratio of the internal EGR gas),i.e., based on the temperature of the internal EGR gas that varies beingaffected thereby. According to this embodiment, therefore, it ispossible to more suitably control the internal combustion engine thanwhen the temperature of gas in the cylinder is not corrected basedthereupon.

In this embodiment, the temperature of gas in the cylinder is calculatedbased on the opening area and the like of the intake valve, and theinternal combustion engine is controlled based on the temperature of gasin the cylinder. In another embodiment, it is also possible to calculatethe temperature of gas in the cylinder based on the opening area and thelike of the exhaust valve, and to control the internal combustion enginebased on the temperature of gas in the cylinder. Namely, this inventioncan be applied not only to the intake valves but also to the exhaustvalves.

Described below is a fourth embodiment of the device for controlling aninternal combustion engine according to the invention. The constitutionof this embodiment is nearly the same as the constitution of the firstembodiment illustrated in FIGS. 1 to 7. The ratio of the internal EGRgas in the cylinder varies as the valve-lifting amount of the intakevalve 2, working angle and opening/closing timing (phase) are varied bythe device 9 for changing the valve-lifting amount and by theopening/closing timing shifting device 11. If the ignition is conductedat a predetermined timing irrespective of a change in the ratio of theinternal EGR gas, an optimum ignition timing is not accomplished, andthe internal combustion engine is not suitably controlled. In order toconduct the ignition at an optimum timing and to suitably control theinternal combustion engine, therefore, the ratio of the internal EGR gasmust be correctly calculated depending upon changes in the valve-liftingamount of the intake valve 2, working angle and opening/closing timing(phase) thereof.

FIG. 36 is a flowchart illustrating a method of calculating the ignitiontiming according to the embodiment. This routine is executed atpredetermined time intervals. When the routine starts as shown in FIG.36, it is, first, judged at step 400 if the engine is being started.When the result is YES, a ratio of the internal EGR gas is correctlycalculated at the starting of the engine, where the amount of the fuelis increased, it is so judged based thereupon that there is no need todetermine the ignition timing, and the routine ends. When the result isNO, on the other hand, the routine proceeds to step 401. At step 401,the normal condition steady-state ratio KEGRb of the internal EGR gas iscalculated based on the valve-lifting amount LT of the intake valve 2,working angle VA, opening/closing timing VT, pressure PM in the intakepipe and the engine rotational speed NE.

FIG. 37 is a diagram illustrating a relationship among the normalcondition steady-state ratio KEGRb of the internal EGR gas, thevalve-lifting amount LT and the opening/closing timing (phase) VT. Asshown in FIG. 37, the normal condition steady-state ratio KEGRb of theinternal EGR gas calculated at step 401 increases with an increase inthe valve-lifting amount LT, or increases as the opening/closing timing(phase) VT advances. FIG. 38 is a diagram illustrating a relationshipamong the normal condition steady-state ratio KEGRb of the internal EGRgas, the working angle VA and the opening/closing timing (phase) VT. Asshown in FIG. 38, the normal condition steady-state ratio KEGRb of theinternal EGR gas calculated at step 401 increases with an increase inthe valve-lifting amount VA.

FIG. 39 is a diagram illustrating a relationship between the normalcondition steady-state ratio KEGRb of the internal EGR gas and thepressure PM in the intake pipe. As shown in FIG. 39, the normalcondition steady-state ratio KEGRb of the internal EGR gas calculated atstep 401 decreases with an increase in pressure PM in the intake pipe.As shown in FIG. 40, the normal condition steady-state ratio KEGRb ofthe internal EGR gas calculated at step 401 decreases with an increasein the engine rotational speed NE.

Reverting to the description of FIG. 36, a steady-state ratio KEGRST ofthe internal EGR gas is calculated (KEGRST←KEGRb×KPA) at step 402 basedon the normal condition steady-state ratio KEGRb of the internal EGR gasand the atmospheric pressure correction coefficient KPA. Namely, thecorrection is effected by taking into consideration the atmosphericpressure having a large ratio of the internal EGR gas. FIG. 41 is adiagram illustrating a relationship between the atmospheric pressurecorrection coefficient KPA and the atmospheric pressure. As shown inFIG. 41, the atmospheric pressure correction coefficient KPA increaseswith an increase in the atmospheric pressure. Namely, the ratio of theinternal EGR gas increases with an increase in the atmospheric pressure.In a modified example of this embodiment, it is also possible tocalculate the correction coefficient based on the back pressure insteadof calculating the correction coefficient KPA based on the atmosphericpressure illustrated in FIG. 41, and to correct the ratio of theinternal EGR gas based on the correction coefficient. FIG. 42 is adiagram illustrating a relationship among the back pressure, the enginerotational speed NE and the intake air amount GN per a revolution. Asshown in FIG. 42, the back pressure increases with an increase in theengine rotational speed NE, or increases with an increase in the intakeair amount GN per a revolution. FIG. 43 is a diagram illustrating arelationship between the back pressure correction coefficient and theback pressure for correcting the ratio of the internal EGR gas. As shownin FIG. 43, the back pressure correction coefficient increases with anincrease in the back pressure. Namely, the ratio of the internal EGR gasincreases with an increase in the back pressure.

In a modified example of the embodiment, further, the steady-state ratioKEGRST of the internal EGR gas can be corrected at step that is notillustrated which is next of step 402 of FIG. 36 based upon the amountof the burnt gas (hereinafter referred to as “amount of blown-back gas”)that is taken in again by the cylinder after being blown back into theintake pipe. FIG. 44 is a diagram illustrating a relationship among theamount of the blown-back gas, the average opening area of the intakevalve 2 (average opening area of the intake valve during the valveoverlapping period) and the average pressure differential before andafter the intake valve 2 (average differential between the pressure inthe cylinder and the pressure in the intake pipe during the valveoverlapping period). As shown in FIG. 44, the amount of the blown-backgas increases with an increase in the opening area of the intake valve2, and increases with an increase in the pressure differential beforeand after the intake valve, i.e., increases as the pressure in thecylinder becomes greater than the pressure in the intake pipe. FIG. 45is a diagram illustrating a relationship between the steady-state ratioKEGRST of the internal EGR gas and the amount of the blown-back gas. Asshown in FIG. 45, the steady-state ratio KEGRST of the internal EGR gasincreases with an increase in the amount of the blown-back gas. Namely,the steady-state ratio KEGRST of the internal EGR gas increases with anincrease in the opening area of the intake valve 2 or increases as thepressure in the cylinder becomes greater than the pressure in the intakepipe. According to this modified example, the ratio of the internal EGRgas is calculated based upon the opening area of the intake valve 2during the valve overlapping period varied by the variable valvemechanism, upon the pressure (pressure in the cylinder) on thedownstream side of the intake valve 2 during the valve overlappingperiod and upon the pressure (pressure in the intake pipe) on theupstream side, and the internal combustion engine is controlled based onthe ratio of the internal EGR gas. Therefore, the internal combustionengine can be suitably controlled by more correctly calculating theratio of the internal EGR gas than when the ratio of the internal EGRgas is calculated based only upon the opening area of the intake valvevaried by the variable valve mechanism or than when the ratio of theinternal EGR gas is not calculated based on the pressure downstream ofthe intake valve during the valve overlapping period or on the pressureon the upstream side.

According to a further modified example, it is allowable to calculatethe instantaneous ratio of the internal EGR gas based upon the openingarea at regular intervals during the valve overlapping period and upon adifference between the pressure in the cylinder (pressure downstream ofthe intake valve 2) during the valve overlapping period and the pressurein the intake pipe (pressure upstream of the intake valve 2) at regularintervals, instead of relying upon the opening area of the intake valveduring the valve overlapping period and the average differential betweenthe pressure in the cylinder and the pressure in the intake pipe duringthe valve overlapping period, in order to control the internalcombustion engine based on the ratio of the internal EGR gas obtained byintegrating the instantaneous ratios of the internal EGR gas. Accordingto this modified example, it is allowed to suitably control the internalcombustion engine by correctly calculating the ratio of the internal EGRgas even when there are great variations in the opening area of theintake valve 2 during the valve overlapping period or in the pressureupstream of the intake valve during the valve overlapping period or inthe pressure downstream thereof.

Reverting to the description of FIG. 36, a change KEGRSM from theprevious time is calculated at step 403 based on the ratio KEGRO of theinternal EGR gas of the previous time and the pressure PM in the intakepipe. FIG. 46 is a diagram illustrating a relationship among the degreeof effect (=1-change KEGRSM from the previous time) due to the ratio ofthe internal EGR gas of the previous time, the ratio KEGRO of theinternal EGR gas of the previous time and the pressure PM in the intakepipe. As shown in FIG. 46, the degree of effect due to the ratio of theinternal EGR gas of the previous time decreases with a decrease in theratio KEGRO of the internal EGR gas of the previous time, and decreaseswith an increase in the pressure PM in the intake pipe. Namely, thechange KEGRSM from the previous time increases with a decrease in theratio KEGRO of the internal EGR gas of the previous time and increaseswith an increase in the pressure PM in the intake pipe.

Reverting to the description of FIG. 36, a ratio KEGR of the internalEGR gas is calculated (KEGR←(KEGRST−KEGRO)×KEGRSM+KEGRO) at step 404based on the steady-state ratio KEGRST of the internal EGR gas, ratioKEGRO of the internal EGR gas of the previous time (=ratio KEGR of theinternal EGR gas calculated at step 404 when the routine was executed inthe previous time) and change KEGRSM from the previous time. Then, atstep 405, an ignition timing SA is calculated based on the ratio KEGR ofthe internal EGR gas, intake air amount GN per a revolution and enginerotational speed NE. As shown in FIG. 47, the ignition timing SAcalculated at step 405 advances with an increase in the ratio KEGR ofthe internal EGR gas, and advances with a decrease in the intake airamount N per one revolution. FIG. 48 is a diagram illustrating arelationship between the ignition timing SA and the engine rotationalspeed NE. As shown in FIG. 48, the ignition timing SA calculated at step405 advances as the engine rotational speed NE increases.

In this embodiment as described above, the ratio of the internal EGR gasis calculated at steps 401 and 404 of FIG. 36 based upon the openingarea of the intake valve 2 that is varied depending upon thevalve-lifting amount LT which is changed by the device 9 for changingthe valve-lifting amount, which is the variable valve mechanism, and theinternal combustion engine is controlled based on the ratio of theinternal EGR gas. According to this embodiment, therefore, it ispossible to suitably control the internal combustion engine by morecorrectly calculating the ratio of the internal EGR gas than when theratio of the internal EGR gas is calculated without considering a changein the opening area of the intake valve 2 by the variable valvemechanism, which is done by the device for controlling internalcombustion engines taught in Japanese Unexamined Patent Publication(Kokai) No. 9-209895. Namely the internal combustion engine can besuitably controlled by correctly calculating the ratio of the internalEGR gas even when the opening area of the intake valve 2 is varied. Morespecifically, the ratio of the internal EGR gas calculated based on theopening area of the intake valve increases with an increase in theopening area of the intake valve 2 as shown in FIG. 37, and the internalcombustion engine is so controlled that the ignition timing SA advancesas the ratio of the internal EGR gas increases as shown in FIG. 47.

In this embodiment, further, the ratio of the internal EGR gas iscalculated at steps 401 and 404 of FIG. 36 based upon the working angleVA of the intake valve 2 that is varied by the device 9 for changing thevalve-lifting amount, which is the variable valve mechanism, and theinternal combustion engine is controlled based on the ratio of theinternal EGR gas. According to this embodiment, therefore, it ispossible to suitably control the internal combustion engine by morecorrectly calculating the ratio of the internal EGR gas than when theratio of the internal EGR gas is calculated without considering a changein the working angle VA of the intake valve 2 due to the variable valvemechanism, which is done by the device for controlling internalcombustion engines taught in Japanese Unexamined Patent Publication(Kokai) No. 9-209895. Namely, the internal combustion engine can besuitably controlled by correctly calculating the ratio of the internalEGR gas even when the working angle VA of the intake valve 2 is varied.More specifically, the ratio of the internal EGR gas calculated based onthe opening area of the intake valve increases with an increase in theworking angle VA of the intake valve 2 as shown in FIG. 38, and theinternal combustion engine is so controlled that the ignition timing SAadvances as the ratio of the internal EGR gas increases as shown in FIG.47.

In this embodiment, further, the ratio of the internal EGR gas iscalculated at steps 401 and 404 of FIG. 36 based upon both the openingarea and the working angle VA of the intake valve 2 that are varied bythe device 9 for changing the valve-lifting amount, which is thevariable valve mechanism, and the internal combustion engine iscontrolled based on the ratio of the internal EGR gas. According to thisembodiment, therefore, it is possible to suitably control the internalcombustion engine by more correctly calculating the ratio of theinternal EGR gas than when the ratio of the internal EGR gas iscalculated based only upon the opening area of the intake valve 2 butnot upon the working angle VA of the intake valve 2 or than when theratio of the internal EGR gas is calculated based only upon the workingangle VA of the intake valve 2 but not upon the opening area of theintake valve 2.

In this embodiment, further, the ratio of the internal EGR gas iscalculated at steps 401 and 404 of FIG. 36 based upon theopening/closing timing (phase) VT of the intake valve 2, pressure PM inthe intake pipe and engine rotational speed NE, and the internalcombustion engine is controlled based on the ratio of the internal EGRgas. According to this embodiment, therefore, it is possible to suitablycontrol the internal combustion engine by more correctly calculating theratio of the internal EGR gas than when the ratio of the internal EGRgas is not calculated based on the opening/closing timing (phase) VT ofthe intake valve 2, pressure PM in the intake pipe and engine rotationalspeed NE.

In this embodiment, further, the ratio of the internal EGR gas iscorrected at step 402 of FIG. 36 based on the atmospheric pressure. Inthe modified example at step 402, the ratio of the internal EGR gas iscorrected based on the pressure in the exhaust pipe, i.e., based on theback pressure, and is further corrected at step 404 based on the ratioKEGRO of the internal EGR gas calculated by the routine in the previoustime. According to this embodiment, therefore, it is possible to moresuitably control the internal combustion engine than when the ratio ofthe internal EGR gas is not corrected based thereupon.

In the above-mentioned embodiment and the modified example, the ratio ofthe internal EGR gas is calculated and the internal combustion engine iscontrolled based thereupon. In place of this, however, it is alsopossible to calculate the amount of the internal EGR gas based on thesame methods as those described above and to control the internalcombustion engine based thereupon. Namely, the tendencies of inclinationof the curves in the above-mentioned diagrams are the same between whenthere is used the ratio of the internal EGR gas and when there is usedthe amount of the EGR gas.

In the above embodiment and the modified example, the ratio or amount ofthe internal EGR gas is calculated based on the opening area of theintake valve, and the internal combustion engine is controlled based onthe ratio or amount of the internal EGR gas. In another embodiment,however, it is also possible to calculate the ratio or amount of theinternal EGR gas based on the opening area of the exhaust valve and tocontrol the internal combustion engine based on the ratio or amount ofthe internal EGR gas. Namely, the invention can be applied not only tothe intake valves but also to the exhaust valves.

Described below is a fifth embodiment of the device for controlling aninternal combustion engine according to the invention. The constitutionof this embodiment is nearly the same as the constitution of the firstembodiment illustrated in FIGS. 1 to 7. The degree of turbulence in thecylinder varies as the valve-lifting amount of the intake valve 2,working angle and opening/closing timing (phase) are varied by thedevice 9 for changing the valve-lifting amount and by theopening/closing timing shifting device 11. If the ignition is conductedat a predetermined timing irrespective of a change in the degree ofturbulence in the cylinder, an optimum ignition timing is notaccomplished, and the internal combustion engine is not suitablycontrolled. In order to conduct the ignition at an optimum timing and tosuitably control the internal combustion engine, therefore, the degreeof turbulence in the cylinder must be correctly calculated dependingupon changes in the valve-lifting amount of the intake valve 2, workingangle and opening/closing timing (phase) thereof.

FIG. 49 is a flowchart illustrating a method of calculating the ignitiontiming according to the embodiment. This routine is executed atpredetermined time intervals. When the routine starts as shown in FIG.49, it is first judged at step 500 if the engine is being started. Whenthe result is YES, the turbulence in the cylinder is correctlycalculated at the start of the engine where the amount of the fuel isincreased, it is so judged based thereupon that there is no need ofdetermining the ignition timing, and the routine ends. When the resultis NO, on the other hand, the routine proceeds to step 501. At step 501,the turbulence CYLTRB is calculated based on the valve-lifting amount LTof the intake valve 2, working angle VA, opening/closing timing VT,pressure PM in the intake pipe and engine rotational speed NE. FIG. 50is a diagram illustrating a relationship among the turbulence CYLTRB inthe cylinder, the valve-lifting amount LT and the opening/closing timing(phase) VT. As shown in FIG. 50, the turbulence CYLTRB in the cylindercalculated at step 501 increases with a decrease in the valve-liftingamount LT, or increases as the opening/closing timing (phase) VT delays.FIG. 51 is a diagram illustrating a relationship among the turbulenceCYLTRB in the cylinder, the working angle VA and the opening/closingtiming (phase) VT. As shown in FIG. 51, the turbulence CYLTRB in thecylinder calculated at step 501 increases with a decrease in the workingangle VA.

FIG. 52 is a diagram illustrating a relationship between the turbulenceCYLTRB in the cylinder and the pressure PM in the intake pipe. As shownin FIG. 52, the turbulence CYLTRB in the cylinder calculated at step 501decreases with an increase in pressure PM in the intake pipe. FIG. 53 isa diagram illustrating a relationship between the turbulence CYLTRB inthe cylinder and the engine rotational speed NE. As shown in FIG. 53,the turbulence CYLTRB in the cylinder calculated at step 501 increaseswith an increase in the engine rotational speed NE.

Reverting to the description of FIG. 49, an ignition timing SA iscalculated based on the turbulence CYLTRB in the cylinder, intake airmount GN per a revolution and engine rotational speed NE. FIG. 54 is adiagram illustrating a relationship among the ignition timing SA, theturbulence CYLTRB in the cylinder and the intake air amount GN per onerevolution. As shown in FIG. 54, the ignition timing SA calculated atstep 502 is delayed with an increase in the turbulence CYLTRB in thecylinder, and is delayed with an increase in the intake air amount GNper a revolution. FIG. 55 is a diagram illustrating a relationshipbetween the ignition timing SA and the engine rotational speed NE. Asshown in FIG. 55, the ignition timing SA calculated at step 502 isadvanced as the engine rotational speed NE increases.

In this embodiment as described above, the turbulence CYLTRB in thecylinder is calculated at step 501 of FIG. 49 based upon the openingarea of the intake valve 2 that is varied depending upon thevalve-lifting amount LT which is changed by the device 9 for changingthe valve-lifting amount, which is the variable valve mechanism, and theinternal combustion engine is controlled based on the turbulence CYLTRBin the cylinder. According to this embodiment, therefore, it is possibleto suitably control the internal combustion engine by more correctlycalculating the turbulence CYLTRB in the cylinder than when theturbulence CYLTRB in the cylinder is calculated without considering achange in the opening area of the intake valve 2 due to the variablevalve mechanism, which is done by the device for controlling internalcombustion engines taught in Japanese Unexamined Patent Publication(Kokai) No. 2000-73800. Namely, the internal combustion engine can besuitably controlled by correctly calculating the turbulence CYLTRB inthe cylinder even when the opening area of the intake valve 2 is varied.More specifically, the turbulence CYLTRB in the cylinder calculatedbased on the opening area of the intake valve increases with an increasein the opening area of the intake valve 2 as shown in FIG. 50, and theinternal combustion engine is so controlled that the ignition timing SAis delayed as the turbulence CYLTRB in the cylinder increases as shownin FIG. 54.

In this embodiment, further, the turbulence CYLTRB in the cylinder iscalculated at step 501 of FIG. 49 based upon the working angle VA of theintake valve 2, opening/closing timing (phase) VT of the intake valve 2,pressure PM in the intake pipe and engine rotational speed NE, and theinternal combustion engine is controlled based on the turbulence CYLTRBin the cylinder. According to this embodiment, therefore, it is possibleto suitably control the internal combustion engine by more correctlycalculating the turbulence CYLTRB in the cylinder than when theturbulence CYLTRB in the cylinder is not calculated based on the workingangle VA of the intake valve 2, opening/closing timing (phase) VT of theintake valve 2, pressure PM in the intake pipe and engine rotationalspeed NE. In the embodiment and in the modified example thereof, theturbulence in the cylinder is calculated based on the opening area ofthe intake valve, and the internal combustion engine is controlled basedon the turbulence in the cylinder. In another embodiment, it is possibleto calculate the turbulence in the cylinder based on the opening areasof the discharge valves and to control the internal combustion enginebased on the turbulence in the cylinder. Namely, the invention can beapplied not only to the intake valves but also to the exhaust valves.

Described below is a sixth embodiment of the device for controlling aninternal combustion engine according to the invention. The constitutionof this embodiment is nearly the same as the constitution of the firstembodiment illustrated in FIGS. 1 to 7 except the points that will bedescribed later. Further, the control routine of this embodiment, whichwill be described later, is executed in combination with the controlroutine of any one of the embodiments described above. In the aboveembodiments, the cam has a nose of a height that is continuouslychanging as shown in FIG. 3. In this embodiment, instead, there areprovided a high-speed cam H having a relatively high cam nose, alow-speed cam L having a relatively low cam noise and anintermediate-speed cam M having a cam nose of a height lyingtherebetween.

FIG. 56 is a flowchart illustrating a method of controlling the camaccording to the embodiment. This routine is executed at regular timeintervals. When the routine starts as shown in FIG. 56, an acceleratoropening degree calculated based on the output value of an acceleratoropening sensor (not shown) is read at step 600. Then, at step 601, anengine rotational speed calculated based on the output value of theengine rotational speed sensor 17 is read. At step 602, a cam isselected based on the accelerator opening degree, engine rotationalspeed and relationship shown in FIG. 57. FIG. 57 is a diagramillustrating a relationship among the accelerator opening degree, theengine rotational speed and the cam to be selected. Referring to FIG.57, when the accelerator opening degree is small and the enginerotational speed is low, the low-speed cam L is selected. The height ofthe cam nose to be selected increases with an increase in theaccelerator opening degree, or the height of the cam nose to be selectedincreases with an increase in the engine rotational speed.

Then, at step 603, it is judged whether it is a timing for changing thecam. When the result is YES, the routine proceeds to step 604 and whenthe result is NO, the routine ends. At step 604, the cam is changed intothe one that is selected. Then, at step 605, a delay in changing the camis estimated based upon the engine rotational speed, upon the coolingwater temperature calculated based on the output value of the coolingwater temperature sensor 20 and upon a relationship illustrated in FIG.58. FIG. 58 is a diagram illustrating the relationship among the delayin changing the cam, the engine rotational speed and the cooling watertemperature. Referring to FIG. 58, the delay in changing the camdecreases with an increase in the engine rotational speed, and decreaseswith an increase in the cooling water temperature.

In a modified example of this embodiment, it is also possible toestimate a delay in changing the cam based on the pressure of theoperation fluid for operating the cam instead of estimating the delay inchanging the cam based on the temperature of the cooling water. FIG. 59is a diagram illustrating a relationship between the delay in changingthe cam and the hydraulic pressure. As shown in FIG. 59, it is estimatedthat the delay in changing the cam decreases with an increase in thehydraulic pressure.

In another embodiment of this embodiment, the delay in changing the camis estimated in advance prior to changing the cam based on the operatingcondition or the hydraulic pressure, and the timing for changing the camis determined by taking the delay into consideration. FIG. 60 is adiagram illustrating a relationship between a moment when an instructionis issued to change the cam and the moment at which the cam is actuallychanged. Referring to FIG. 60, a delay in changing the cam (=timet1-time t0) is estimated and when it is attempted to change the cam atthe time t1, an instruction to change the cam is issued at the time t0.

FIG. 61 is a flowchart illustrating a method of calculating a fuelinjection amount according to the embodiment. This routine is executedat regular time intervals. When the routine starts as shown in FIG. 61,first, an output value of the air flow meter 19 is read at step 700.Then, at step 701, an engine rotational speed calculated based on theoutput value of the engine rotational speed sensor 17 is read. Then, anestimated cam selection value is read at step 702. That is, a flagrepresenting a cam selected at step 602 of FIG. 56 is read. Then, atstep 703, an intake air amount per a revolution is calculated by thesame method as the one of the above-mentioned embodiment. At step 704, aresponse correction coefficient is calculated based on the type of thecam, the engine rotational speed, the intake air amount per a revolutionand a relationship shown in FIG. 62. FIG. 62 is one diagram illustratingthe relationship among the response correction coefficient, type of thecam, engine rotational speed and intake air amount GN per onerevolution. Then, at step 705, a fuel injection amount is calculatedbased on the intake air amount per a revolution and a relationship shownin FIG. 63. FIG. 63 is a diagram illustrating the relationship betweenthe fuel injection amount and the intake air amount per a revolution.

FIG. 64 is a flowchart illustrating a routine for calculating anignition timing according to the embodiment. This routine is executed atregular time intervals. When the routine starts as shown in FIG. 64,first, an intake air amount per one revolution is read at step 800.Then, at step 801, the engine rotational speed is read. Next, anestimated cam selection value is read at step 802. Thereafter, at step803, an ignition timing is calculated based on the type of the cam, theengine rotational speed, the intake air amount per one revolution and arelationship shown in FIG. 65. FIG. 65 is a diagram illustrating therelationship among the ignition timing, type of the cam, enginerotational speed and intake air amount GN per one revolution.

In a modified example of the above embodiment, it is also possible, asrequired, to employ the intake and exhaust valves driven by theelectromagnetic force or the hydraulic pressure instead of using theintake and exhaust valves driven by the cams.

According to the present invention as described above, the internalcombustion engine can be controlled based not only upon the peakcombustion pressure in the cylinder but also upon a pressure in thecylinder at a moment other than the peak combustion pressure. Namely,the internal combustion engine can be suitably controlled even when theopening areas or the working angles of the intake and exhaust valves arevaried.

Further, the invention not only controls the internal combustion enginebased simply on the cylinder wall temperature but also controls theinternal combustion engine based on a correctly measured temperature ofgas in the cylinder. The invention, further, makes it possible tosuitably control the internal combustion engine even when the openingareas or the working angles of the intake and exhaust valves are varied.

According to the invention, further, the amount of the internal EGR gasis correctly calculated even when the opening areas or the workingangles of the intake and exhaust valves are varied, and the internalcombustion engine is suitably controlled based on the calculated amountof the internal EGR gas.

Lastly, according to the invention, the degree of turbulence in thecylinder is correctly estimated even when the opening areas of theintake vales are varied by the variable valve mechanism, and theinternal combustion engine is suitably controlled based on the degree ofturbulence in the cylinder.

1. A device for controlling an internal combustion engine, comprising a variable valve mechanism for at least either the intake valves or the exhaust valves thereby to control the internal combustion engine based on a degree of turbulence in the cylinder that is estimated based on the opening area of the intake valve varied by the variable valve mechanism, wherein it is so estimated that the degree of turbulence in the cylinder increases with a decrease in the opening area of the intake valve varied by the variable valve mechanism, and the internal combustion engine is controlled based on the estimated degree of turbulence in the cylinder.
 2. A device for controlling an internal combustion engine according to claim 1, wherein a degree of turbulence in the cylinder is estimated based upon the working angle of the intake valve, phase of the intake valve, pressure in the intake pipe and engine rotational speed in addition to said opening area, and the internal combustion engine is controlled based upon the estimated degree of turbulence in the cylinder.
 3. A device for controlling an internal combustion engine according to claim 1, wherein it is so estimated that the degree of turbulence in the cylinder increases with a decrease in the working angle of the intake valve, and the internal combustion engine is controlled based on the estimated degree of turbulence in the cylinder.
 4. A device for controlling an internal combustion engine according to claim 1, wherein an ignition timing of the internal combustion engine is controlled based upon the estimated degree of turbulence in the cylinder, and the ignition timing delays with an increase in the estimated degree of turbulence in the cylinder. 